1. Field of the Invention
The present invention relates to differentials, and more particularly, to traction enhancing differentials having cone clutch elements.
2. Description of the Related Art
Differentials are well known in the prior art and allow a pair of output shafts operatively coupled to an input shaft to rotate at different speeds, thereby allowing the wheel associated with each output shaft to maintain traction with the road while the vehicle is turning. Such a device essentially distributes the torque provided by the input shaft between the output shafts. However, the necessity for a differential which limits the differential rotation between the output shafts to provide traction on slippery surfaces is well known.
The completely open differential, i.e., a differential without clutches or springs, is unsuitable in slippery conditions where one wheel experiences a much lower coefficient of friction than the other wheel, for instance, when one wheel of a vehicle is located on a patch of ice and the other wheel is on dry pavement. In such a condition, the wheel experiencing the lower coefficient of friction loses traction and a small amount of torque to that wheel will cause a "spin out" of that wheel. Since the maximum amount of torque which can be developed on the wheel with traction is equal to torque on the wheel without traction, i.e. the slipping wheel, the engine is unable to develop any torque and the wheel with traction is unable to rotate. A number of methods have been developed to limit wheel slippage under such conditions.
Prior methods of limiting slippage between the side gears and the differential casing include use of a frictional clutch mechanism having a frusto-conical engagement structure and a bias mechanism, usually spring loaded, to apply an initial preload between the side gears and the differential casing. By using a frictional clutch with an initial preload a minimum amount of torque can always be applied to the wheel having traction, i.e. the wheel located on dry pavement. The initial torque generates gear separating forces which further engage the frictional clutch and develop additional torque.
The initial preload initiates the development of side gear separating forces which provide further braking action between the side gears and the differential casing. In general, gear separating forces are forces induced, due to the angle of contact or "pressure angle", on any set of meshing gears by the application of torque to the gears and which tend to separate the gears. In a differential, the development of torque will create side gear separating forces which tend to move the side gears away from the pinion gears. When one wheel is on a surface having a low coefficient of friction, the initial preload creates some contact and frictional engagement between the differential casing and the clutch mechanism disposed between the side gears and the differential casing to allow the engine to provide torque to the wheel having traction. This initial torque transfer induces gear separating forces on the side gears which tend to separate the side gears to further frictionally engage the clutch mechanism with the casing. The increased frictional engagement of the clutch allows more torque to be developed between the casing and the clutch element, thus further increasing the side gear separating forces and limiting the slippage between the side gears and the differential casing.
It is well known in the art to use frusto-conical clutch elements providing, on the outside surfaces thereof, a spiral structure which winds helically about the clutch element from its base to its tip, the tip comprising the annular edge resulting when the right circular cone is truncated at a plane parallel to its base, producing a frustum. The spiral structure provides a clutch engagement surface which frictionally engages an adjacent frusto-conical clutch interior surface of the differential casing. Generally, such cones are provided with a double helical structure, each helix beginning and ending at radially opposite points of the outside surface of the clutch element. These frusto-conical clutch elements are made of generally ferrous material and are produced using powdered metal or machined casting methods.
FIG. 1 illustrates one embodiment of prior art limited slip differential 10 having helical cone clutch elements. Differential 10 comprises casing 20, which includes casing parts 22 and 24 which are assembled via threaded joint 26. Casing part 22 includes radial flange 28, to which a ring gear (not shown) is attached by, for example, threaded fasteners (not shown). Torque output from a vehicle transmission applied to the ring gear causes differential casing 20 to rotate about axis 29. Casing parts 22 and 24 are provided with hollowed hub portions 30 and 32, respectively, through which extend output shafts or axles 34 and 36 along longitudinal axis 29. Fixed for rotation with the end of each axle 34 and 36 in the interior of casing 20 are bevel side gears 38 and 40, respectively. In the embodiment shown in FIG. 1, each side gear 38 and 40 is fixed for rotation with cone clutch element 42 and 44, respectively, having the above mentioned double helical structure about their outside surfaces. Cone clutch elements 42 and 44 are usually identical and do not necessarily provide helical structures which spiral outwardly from the center of case 20 along axis 29 as mirror images of one another. Notably, other embodiments of limited slip differentials may provide only one cone clutch member.
Intermeshed with the teeth of side gears 38, 40 are pinion gears 46, 48. The pinion gears rotate about cross shaft 50 which extends therethrough and is attached to casing 20 to rotate therewith. Thus pinion gears 46, 48 rotate about cross shaft 50 and revolve about axis 29 with casing 20. Cross shaft 50 is usually retained to casing 20 with a fastener such as bolt 52. Disposed between the facing surfaces of bevel side gears 38, 40 is some form of preload mechanism 54. In the shown embodiment preload mechanism 54 comprises a plurality of compression springs 56 and bearing plates 58, 60. Bearing plates 58 and 60 bear on the facing surfaces of bevel side gears 38 and 40, respectively, urging them apart under the influence of springs 56. This separating force is imparted through the side gears to the cone clutch elements 42, 44, urging their outside frusto-conical surfaces into relatively light frictional engagement with mating frusto-conical clutch seat surfaces 62, 64 of the interior of casing part 22. When the wheels (not shown) attached to axles 34, 36 have equal traction, input torque to casing flange 28 is distributed approximately equally therebetween, transmitted from casing 20 to cross pin 50, to pinion gears 46, 48, to side gears 38, 40 and then to axles 34, 36, which generally rotate at the same speed as casing 20. Under this condition, little appreciable torque is transmitted directly from casing 20 to side gears 38, 40 and axles 34, 36 through cone clutch elements 42, 44 because the frictional engagement between clutch seat surfaces 62, 64 and cone clutch elements 42, 44 is generally rather light and minor clutch slippage is allowed when turning. However, as one of the wheels attached to axles 34, 36 loses traction, the two axles and the cone clutch elements fixed to rotate therewith begin to rotate at different speeds relative to each other and to rotating casing 20. Under this condition, separation forces acting between pinion gears 46, 48 and side gears 38, 40, plus the spring preload forces, in conjunction with the sliding relative motion between clutch elements 42, 44 and seat surfaces 62, 64, cause frictional torque transfer between cone clutch elements 42, 44 and casing surfaces 62, 64, braking the axle rotating faster than casing 20 and transferring torque from casing 20 to the slower moving axle.
FIG. 2 shows a typical embodiment of prior art frusto-conical clutch element 42 (assumed identical to element 44) having, on its outside surface, helical structures providing two rather narrow, spiraling clutch engagement surfaces 66, 68 which coincide with a conic surface defined by imaginary element lines (two of which are represented by reference numerals 71 and 72) extending from base 78 of a right circular cone to its vertex 128 (FIG. 13). Each clutch engagement surface 66, 68 is bounded by a pair of spiraling, parallel lateral edges 69, 70. The prior art double helix cone clutch element of FIGS. 2 and 3 has spiraling clutch engagement surfaces 66, 68 begin at approximately radially opposite sides of base 78 and end at approximately radially opposite sides of tip 80. Each surface 66, 68 spirals approximately 360.degree. circumferentially about the frusto-conical shape of element 42, with the full width of each surface 66, 68, i.e., the perpendicular distance between parallel lateral edges 69, 70, exposed over a circumferentially spiraling angle greater than 180.degree. but less than 360.degree..
Lines 71 and 72, located on radially opposite sides of element 42, are typically separated by included angle .theta. ranging from about 10.degree. to about 25.degree., depending on performance characteristics, with the smaller angle providing more aggressive clutch performance and the larger angle providing less aggressive clutch performance. It has been found that if the cone clutch angle is too small, there may be difficulty in releasing the clutch due to the wedging effect between the interengaging clutch surfaces and, if the cone clutch angle is too large, excessive preload pressure will be required to prevent slippage. Both surfaces 66, 68 of element 42 frictionally engage, in operation, mating clutch seat surface 62 in the interior of rotating casing part 22 (FIG. 1). Further, the helical structure of the outside surface of the shown clutch element provides two spiraling grooves or reliefs 74, 76 for channeling oil to and from the interfacing clutch surfaces. Examples of limited slip differentials employing such helically surfaced cone clutch elements are disclosed in U.S. Pat. Nos. 4,612,825 (Engle), 5,226,861 (Engle), 5,556,344 (Fox), and U.S. patent application Ser. No. 09/030,602 (Forrest et al), filed Feb. 25, 1998, which are assigned to the assignee of the present invention and expressly incorporated herein by reference.
A limited slip differential's ability to transfer the torque which is applied to the rotatable casing to the axle shafts is characterized by its bias ratio (BR), which is defined as the ratio of the torque applied to the higher torque axle (T.sub.high), i.e., the relatively slower spinning or nonrotating axle, divided by the torque applied to the lower torque axle (T.sub.low), i.e., the faster spinning axle. The total torque (T.sub.tot) transferrable from the rotating differential case to the axles equals the sum of the torque applied, equally or unequally, to each of the two axles. Thus, the bias ratio can be expressed as follows: EQU BR=T.sub.high /T.sub.low =(T.sub.tot -T.sub.low)/T.sub.low (Equation 1)
A higher bias ratio means that the two axles attached to the differential act more like a solid axle in that the differential is better able to transfer torque applied to the rotating casing unequally to each axle, for in a limited slip differential, the torque applied to the rotating casing tends to be transferred to each axle depending on the traction available at that axle's wheel. In contrast, an open differential, i.e., a differential having no limited slip feature, which has a theoretical bias ratio of 1:1, transfers the torque applied to the rotating casing to each axle equally, regardless of the traction available at each wheel.
With reference now to FIG. 4, a linear graph which utilizes typical values of torque along its left hand vertical and horizontal scales, straight line 82 represents the theoretical relationship between the torque applied to the loose wheel or lower torque (faster spinning) axle (T.sub.low) on the horizontal scale and T.sub.tot, for an open differential with the opposite axle fixed so as not to rotate, the "wheel" of the nonrotating axle thus having maximum traction. The theoretical bias ratio for an open differential being 1:1, each point on line 82 has a value of T.sub.tot, which is twice that of T.sub.low.
Straight line 84 represents the theoretical relationship between T.sub.low and T.sub.tot under the assumption that T.sub.high is held at a maximum value of 12,500 inch pounds, i.e., one half the 25,000 inch pounds total torque applied to casing 20. At each point on line 84, T.sub.tot equals T.sub.low plus 12,500 inch pounds (T.sub.high). Line 84 thus represents a solid axle condition. A limited slip differential cannot transfer less torque than an open differential nor more torque than a solid axle. Therefore, between lines 82 and 84 is the total operating envelope for differentials, comprising a range of T.sub.low and T.sub.tot relationships for limited slip differentials which may be plotted linearly along lines stemming from the intersection of the ordinate and the abscissa, each such linear line representing a different bias ratio. For example, bias line 86 represents a bias ratio of 1.67:1 and bias line 88 represents a bias ratio of 2.5:1. It can be seen, therefore, that the slope of the bias line is determined by clutch effectiveness.
The operation of a theoretical limited slip differential will now be further explained with reference to FIG. 4: Bias line 86, representing a BR of 1.67:1, extends from point 0,0, at the intersection of the ordinate and the abscissa, to point A, where it intersects line 84. (It should be noted that if a clutch biasing mechanism such as preload mechanism 54 (FIG. 1) were provided in the example differential, straight line 86 would intersect the ordinate at a value higher than zero.) Along line 86, to the left of point A, the differential controls the distribution of torque T.sub.tot from the rotating casing to the non-rotating, "tight" wheel axle and the spinning, "loose" wheel axle. The amount of torque T.sub.high which may be applied to the tight wheel axle is limited to the maximum traction available to that axle's wheel which, in this example, is 12,500 pound inches. The loose wheel traction is, during testing, variably controlled by means of a brake to set the amount of loose wheel torque T.sub.low.
At all points along line 86, the tight wheel axle has more traction available to it than is utilized, and the differential governs how much of the total torque T.sub.tot is transferred from the casing to the axles; in other words, the differential is still differentiating. For example, although the tight wheel has 12,500 pound inches of traction available to it, equation 1 can be used to reveal that at the point on line 86 where T.sub.tot is 10,000 pound inches and T.sub.low is 3750 pound inches, T.sub.high is only 6250 pound inches. At point A, where T.sub.tot is 20,000 pound inches and T.sub.low is 7500 pound inches, T.sub.high reaches the maximum traction level of the tight wheel of 12,500 pound inches.
As T.sub.low increases above 7500 pound inches, and T.sub.high exceeds the maximum traction level available to the tight wheel, the differential no longer controls the total amount of torque T.sub.tot transferred from the rotating casing to the axles, and bias curve 86 no longer applies; in other words, the differential stops differentiating. At values of T.sub.low beyond 7500 pound inches, both the tight and loose wheels spin, and relationship between T.sub.tot and T.sub.low follows curve 84 from point A onwards, simulating a solid axle in that each unit increment of T.sub.low is correspondingly added to T.sub.tot. Thus, at points on line 84 to the right of point A, the amount of total torque T.sub.tot transferred from the casing to the axles depends solely on the amount of traction available to the loose wheel.
Because preloaded clutches are usually always engaged, they are susceptible to wear. And although frusto-conical, helical clutch elements as described above are initially effective in providing adequate clutched engagement, over repeated use that effectiveness degrades significantly, reducing the amount of torque which can be transferred between cone clutch element 42 and differential casing 20.
Referring again to FIG. 4, line 90 represents measured values of T.sub.tot for given T.sub.low values for limited slip differential 10 as shown in FIG. 1, having two double helical cone clutches as described above. The total area of clutch engagement surfaces 66, 68 is about 5.2 square inches per clutch element and the included angle .theta. between element lines 71, 72 is 25.degree.. The data generating line 90 taken from a particular differential unit prior to durability testing. This same differential unit, after having undergone 400 miles of simulated highway driving with a normal sized tire and wheel attached to one axle and a mini-spare tire attached to the other, a condition designed to induce clutch slippage and thus burnish the clutch surfaces, produced the measured values of T.sub.tot for given T.sub.low represented by line 92. As can be seen, the bias ratio of this differential unit degraded appreciably through use. The right hand side of FIG. 4 has been scaled vertically to compare directly, through lines 94 and 96, the relationship between bias ratio and T.sub.low before and after durability testing, respectively. A comparison of lines 94 and 96 illustrates a substantial degradation in clutch performance over the operating range of T.sub.low values after the clutch surfaces have worn. This resulting reduction in the bias ratio may lead to undesirable repair and replacement costs. A limited slip differential with improved bias ratio durability is thus desirable, particularly if this improvement can be accomplished without substantially increasing variable cost, package size or weight. Thus, it is desired to provide this advantage with a cone clutch element which is directly interchangeable with element 42.